In rotary rolling piston hermetic compressors with low back pressure and high internal pressure in the shell, it occurs the phenomenon of the passage (penetration) of the lubricating oil to the interior of the cylinder, into the suction and the discharge chambers.
The oil contained at the bottom of the shell and submitted to the high gas pressure inside the shell, will be elevated by an oil pump or another device, until it reaches the crankshaft, from which the oil is then radially displaced through the gaps between the annular end faces of the rolling piston and the bearing covers, entering the cylinder internal chambers.
This penetration of oil at a high temperature into the cylinder causes on the functioning and performance of the compressor the effects discussed hereinafter.
On penetrating into the suction chamber, the oil, at a high temperature, warms up the incoming suction gas, causing an increase of its specific volume and, therefore, reducing the suction chamber filling up capacity. Thus, the gas mass which fills the cylinder suction chamber is reduced by the effect of the increase of the gas specific volume. Besides this inconvenience, it should be noted that the oil volume itself which penetrates in the suction chamber takes gas filling space; however, this effect is of a quite secondary importance in relation to the heating effect.
The above mentioned problem causes a decrease in the compressor pumping capacity as a result of the lubricating oil penetration into the cylinder.
In turn, on penetrating into the cylinder compression chamber, the oil will be, during a great part of the compression period, at a higher temperature than the gas temperature under compression, also causing the heating of such gas and increasing its specific volume. This phenomenon results in an increase in the work required for compressing the gas and, consequently, an increase in the compressor energy consumption. This fact can be verified in FIG. 2 of the accompanying drawings, where a pressure X angle of rotation diagram is presented for demonstrating that the compression pressure raises faster when the oil leakage to the interior of the compression chamber increases as represented by the dashed line Q.
These two effects when combined, contribute to considerably drop the compressor volumetric and energy efficiency.
On the other hand, the presence of the lubricating oil flow carries out two favorable functions which are fundamental to the compressor functioning.
The first one, and the most obvious one, is the lubrication of the movable parts involved.
The second one is the sealing of all clearances between the movable parts, thus avoiding the gas direct leakage from the interior of the cylinder to the interior of the shell, which leakage, in case of happening, can be even more prejudicial to the compressor, in terms of capacity drop, than the gas overheating by the oil.
This property of the oil of sealing gaps between the movable parts, acts on the cylinder internal leakages (from the compressor chamber to the suction chamber at low pressure), and on the leakages from the compression chamber to the interior side of the shell.
In the more specific case of the oil which penetrates radially into the cylinder through the rolling piston end faces, the lubricating oil prevents the gas from leaking from the compression chamber to the crankshaft internal parts and from the latter to the interior of the shell.
Therefore, the amount of oil which gets into the cylinder must be controlled at an optimum level, i.e., at a minimum level in order to make possible to have the sealing of the gas leakages and, at the same time, only a minimum gas heating in the interior of the cylinder.
A well-known way to control the amount of oil that penetrates into the cylinder, by the gaps of the rolling piston end faces, is to reduce such gaps up to a minimum level in which the losses by friction between the rolling piston end faces and the bearing cover faces do not reach such a value to completely cancel the gains resultant from the oil flow reduction through said gaps.
Despite the possibility of reducing the rolling piston end gaps in a way to get an advantageous reduction in the amount of lubricating oil which penetrates into the cylinder, the obtained gain, in terms of energetic efficiency of the compressor, will always be inferior to what would be reached by the exclusive reduction of the oil flow, due to greater friction loss as a result of minor or greater reduction of the said gaps.
In the development of the present invention, it was found out that an increase in the radial path of the lubricant oil flow through the opposite axial gaps of the rolling piston makes it possible to reduce, by at least 10%, the oil flow to the interior of the suction and discharge chambers, without substantially increasing the friction losses between the movable parts.
From the equation that models the radial flow of oil through the rolling piston end faces (viscous flow between parallel discs) it can be noted that the flow is controlled by the gap (.delta.R) and by the thickness of the rolling piston wall or, more precisely, by the relation: ##EQU1##
The behavior of such function can be observed in the graph of FIG. 3. The rolling piston dimensions found in the presently marketed compressors with small displacement volume present an external diameter (ext. .sup.100)/internal diameter (int. .sup..phi.) relation between 1.40 and 1.55, defining rolling pistons of thin walls.
The invention aims to define as rolling pistons of thick wall those which present the external diameter/internal diameter relation.gtoreq.1.63, approximately.
It can be noticed by the graph of FIG. 3 that, below about the value of external diameter/internal diameter=1.63, the slope of curve 1n.sup.-1 external diameter/internal diameter is quite accentuated, becoming gradually less steep after said value has reached about 1.6.
It should be remembered that the behavior of the curve that represents the oil flow toward the interior of the cylinder is proportionally reflected in the compressor performance already explained, i.e., the greater the value of the function 1n.sup.-1 (external diameter/internal diameter), the greater will be the oil flow and worse the volumetric and energy efficiency of the compressor. Therefore, it is important to notice that the penetration of oil inward the cylinder can have a reduction of at least 10% with a dimensioning of the rolling piston diameters in a way to get a relation (external diameter/internal diameter) from 1.63 on, regarding the range commonly used, the rolling piston being dimensioned in order to get a relation of external diameter/internal diameter.gtoreq.1.63, approximately.
It is also important to mention that diameter relations of 1.63 on, up to nearly 2.22 are perfectly feasible in the production processes normally used for rotary rolling piston compressors and yet, there is no impediment or disadvantage in terms of the compressor performance when using such relations.
One possible disadvantage would be the increase of friction losses between the piston faces and the cylindrical chamber end walls due to the increase of the contact surface. However, the increase of the friction losses does not effectively occur because, with the increase of the contact surface, there is a tendency to have a reduction of the angular velocity of the piston over its own shaft which compensates the loss. Besides, such loss by friction is of the order of magnitude at least one time smaller than the losses caused by the heated oil, when the gap .delta.R is the usual one. Therefore, such loss could be neglected anyway.
It was also observed that certain prior art rotary hermetic compressors with high displacement volume (higher than about 7 cc) and/or low internal pressure in the shell (low side compressor) present external diameter/internal diameter relations for the rolling piston within the range of 1.63 to 2.22. However, the existance of rotary hermetic compressors with high displacement volume and/or low internal pressure in the shell having such dimensional relation is merely casual. There is not any technical literature suggesting the use of said dimensional relation to obtain a reduction of the oil flow to the interior of the suction and discharge chambers.
According to the available technical literature, it can be affirmed that the fact of existing rotary hermetic compressors presenting said dimensioning relation is a simple consequence of the fact that the displacement volume, which was designed to correspond to a preset capacity, is high (higher than 7.1 cc), thus making high the values of the cylinder and rolling piston radii, whereas the shaft radius is determined by its minimum possible value, due to the strength of the material which is used. In other words, it can be said that in hermetic compressors with high displacement volume, the shaft radius is small enough to allow a relatively small radius for the eccentric and, consequently, an also small internal radius for the rolling piston. Thus, the external diameter/internal diameter relation of the rolling piston can be situated within the above mentioned range only casually.
Although there are rotary hermetic compressors with a rolling piston presenting a thick wall, it should be noted that such compressors are of the "low-side" type (low internal pressure in the shell). Nevertheless, there are fundamental differences regarding the finality and functioning of a thick rolling piston in a "high-side" (high internal pressure in the shell) hermetic compressor and in a "low-side" hermetic compressor (low internal pressure in the shell). In the low-side compressor, the low internal pressure in the shell does not act on the oil which, therefore, it is not allowed to reach the interior of the cylinder through the gaps between the movable parts, as it occurs in the high-side compressor. Thus, in the low-side compressor, the oil does not act as a sealant against the gas leakages through the gaps, the compressed gas in the compression chamber tending to leak through the gaps between the movable parts, more specifically between the rolling piston end faces and the bearing covers. The flow in said gaps is, therefore, of gas leakage in the low-side compressor, and not of oil penetration as it occurs in the high-side compressors.
Reducing the gaps or increasing the rolling piston thickness in the low-side compressors has the finality of avoiding the gas leakage in the compression chamber, and not of controlling the problem of oil flow to the interior of the cylinder, as it occurs in the high-side compressors. Thus, the finality of increasing the thickness of the rolling piston in both types of compressors is completely different.
As the internal diameter of the rolling piston is approximately the same as the diameter of the eccentric portion of the crankshaft, the desired relation can be represented as follows (see FIG. 4B): ##EQU2## where: RR=external radius of the rolling piston
Rr=internal radius of the rolling piston PA0 Re=radius of the crankshaft eccentric portion PA0 Ec=eccentricity of the eccentric portion PA0 Rs=radius of the compressor shaft PA0 Re=radius of the eccentric portion PA0 RR=external radius of the rolling piston PA0 Rr=internal radius of the rolling piston PA0 Rm=radius of the crankshaft end portion PA0 Rs=radius of the crankshaft
As the external radius RR of the rolling piston is determined in relation to the cylinder diameter that is designed for the compressor, the relation (1) above can be achieved by changing the values of the piston internal radius Rr and, consequently, the radius Re of the crankshaft eccentric portion.
In the known rotary hermetic compressors (having a displacement volume above 7 cc), presenting the dimensional relation (1) above, the internal radius Rr of the rolling piston (or radius Re of the crankshaft eccentric portion) is generally large enough to allow the following dimensional relation: EQU Rr.apprxeq.Re=Ec+Rs (2)
where:
The dimensional relation (2) above is shown in FIG. 4A, though this prior art solution does not necessarily present the dimensional relation (1) simultaneously.
When the compressor presents the dimensional relation (2) above, the radius Rm of the shaft end portion can be maintained equal to the radius Rs of the crankshaft, i.e., Rm=Rs, without causing any problem of assembling the rolling piston on the eccentric portion of the crankshaft, as illustrated in FIG. 4A, where the contour of the eccentric portion is tangent to the crankshaft remainder contour.
Nevertheless, in the rotary hermetic compressors with small displacement volume (less than 7 cc) and high internal pressure in the shell, the reduction of the internal radius Rr of the piston (or radius Re of the eccentric portion), in order to achieve a radial extension of the piston wall within the relation (1), can make it impossible to have, due to the eccentricity Ec required in the compressor design and to the minimum diameter required for the shaft, both dimensional relations (1) and (2) simultaneously. In these prior art compressors, the dimensional relation (1) can only be obtained in conjunction with the following dimensional relation (FIG. 4b): EQU 2Rr&lt;Rm+Ec+Re (3)
In this situation, the contour of the crankshaft eccentric portion is not tangent to the crankshaft contour anymore, becoming secant to the latter, avoiding that the rolling piston be mounted at the crankshaft eccentric portion.